Hermetic type compressor and refrigeration cycle apparatus

ABSTRACT

As a hermetic type compressor, a motor portion and a compression mechanism portion that are coupled to the motor portion with a rotating shaft interposed therebetween are accommodated in a closed vessel. The compression mechanism portion comprises a cylinder that comprises an internal diameter hole, and a main bearing and a sub-bearing in which a bearing hole that journals the rotating shaft is provided and the internal diameter hole of the cylinder is closed to form a compression chamber in the compression mechanism portion. The main bearing and the sub-bearing have a circular groove that is opened toward the compression chamber side, an inner circumferential surface of the circular groove is tapered such that a diameter increases gradually from the compression chamber side toward an opposite side of the compression chamber side, and a depth of the circular groove is set to 40% of a diameter of the bearing hole.

CROSS REFERENCE TO RELATED APPLICATIONS

This is a Continuation Application of PCT Application No.PCT/JP2009/059719, filed May 27, 2009, which was published under PCTArticle 21(2) in Japanese.

This application is based upon and claims the benefit of priority fromprior Japanese Patent Application No. 2008-139682, filed May 28, 2008,the entire contents of which are incorporated herein by reference.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a hermetic type compressor whosebearing structure is improved and a refrigeration cycle apparatus thatincludes the hermetic type compressor to form a refrigeration cycle.

2. Description of the Related Art

Frequently, a rotary hermetic type compressor is used in therefrigeration cycle apparatus. In the rotary hermetic type compressor, amotor portion and a compression mechanism portion that is coupled to themotor portion via a rotating shaft (crankshaft) interposed therebetweenare accommodated in a closed vessel. In this kind of compressor, arefrigerant is introduced into a compression chamber formed in acylinder and compressed, whereby a compressive load acts on the rotatingshaft.

Accordingly, the rotating shaft generates a flexural deformation, and arotating shaft portion in a flexure direction and a bearing thatjournals the rotating shaft come into partial contact with each otherunless some sort of measure is taken. Smooth rotation of the rotatingshaft is spoiled, which leads to damage of the rotating shaft andbearing. Therefore, for example, Jpn. Pat. Appln. KOKAI Publication No.2004-124834 proposes a bearing structure in order to properly bear theflexural deformation of the rotating shaft.

In the technique proposed in Jpn. Pat. Appln. KOKAI Publication No.2004-124834, according to the flexural deformation of the rotating shaftdue to the compressive load in the cylinder, a groove is provided on acylinder side of a main bearing to allow the flexural deformation of themain bearing, and a center of an internal diameter on the motor side ofthe main bearing is eccentrically disposed by a predetermined amountwith respect to a center of an internal diameter on the cylinder side ina direction of the flexural deformation of the rotating shaft.

BRIEF SUMMARY OF THE INVENTION

However, in the groove on the cylinder side of the main bearing in thetechnique, a diameter of an inner circumferential surface of the mainbearing is kept constant over a total length, and a thickness betweenthe inner circumferential surface of the groove and an innercircumference of a bearing hole is also kept constant over the totallength.

Accordingly, although the partially strong contact between the rotatingshaft and the bearing can be avoided in a certain range of the groove bythe flexure of the bearing, rigidity of the bearing increases rapidly atan end of the groove, and a contact load is concentrated on the end ofthe groove. Therefore, local abrasion is generated, and bearingreliability cannot sufficiently be enhanced.

In view of the foregoing, an object of the invention is to provide ahermetic type compressor in which, according to the flexural deformationof the rotating shaft due to the compressive load in the cylinder,uneven contact with the rotating shaft is prevented in at least one ofthe main bearing and sub-bearing, thereby achieving the enhancement ofthe reliability and a longer operation life.

Another object of the invention is to provide a refrigeration cycleapparatus that includes the hermetic type compressor to form therefrigeration cycle, thereby improving refrigeration efficiency.

A hermetic type compressor of the present invention comprises, a motorportion and a compression mechanism portion that are coupled to themotor portion with a rotating shaft interposed therebetween areaccommodated in a closed vessel, the compression mechanism portioncomprises a cylinder that comprises an internal diameter hole; and amain bearing and a sub-bearing in which a bearing hole that journals therotating shaft is provided and the internal diameter hole of thecylinder is closed to form a compression chamber in the compressionmechanism portion, at least one of the main bearing and the sub-bearinghave a circular groove that is opened toward the compression chamberside, an inner circumferential surface of the circular groove is taperedsuch that a diameter increases gradually from the compression chamberside toward an opposite side of the compression chamber side, and adepth of the circular groove is set to at least 40% of a diameter of thebearing hole.

A refrigeration cycle apparatus of the present invention comprises, thehermetic type compressor; a condenser; an expansion device; and anevaporator.

BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWING

FIG. 1 is a refrigeration cycle configuration diagram of a refrigerationcycle apparatus according to a first embodiment of the invention and alongitudinal sectional view of a hermetic type compressor.

FIG. 2 is an enlarged longitudinal sectional view of a compressionmechanism portion of the hermetic type compressor.

FIG. 3 is an enlarged longitudinal sectional view of a compressionmechanism portion of a hermetic type compressor according to a secondembodiment of the invention.

FIG. 4 is a longitudinal sectional view of a main part of a hermetictype compressor according to a third embodiment of the invention.

FIG. 5 is a longitudinal sectional view of a main part of a hermetictype compressor according to a fourth embodiment of the invention.

FIG. 6 is a characteristic diagram of a circular groove depth effect inthe invention.

FIG. 7 is a characteristic diagram of a circular groove minimum wallthickness effect in the invention.

FIG. 8 is a characteristic diagram of a circular groove minimum sealwidth effect in the invention.

FIG. 9 is a characteristic diagram of a circular groove slope effect inthe invention.

FIG. 10 is a longitudinal sectional view of a hermetic type compressoraccording to a modification of the third embodiment of the invention.

FIG. 11 is a plan view of a discharge valve mechanism mounted on anintermediate partition plate of the modification.

FIG. 12 is a sectional view of an intermediate partition plate and adischarge valve mechanism of a first example of the modification.

FIG. 13 is a sectional view of an intermediate partition plate and adischarge valve mechanism of a first example of the modification.

DETAILED DESCRIPTION OF THE INVENTION

Embodiments of the invention will be described below with reference tothe drawings. FIG. 1 is a longitudinal sectional view of a hermetic typecompressor 1 and a refrigeration cycle configuration diagram of arefrigeration cycle apparatus R.

In FIG. 1, the numeral 1 designates a hermetic type rotary compressor(hereinafter simply referred to as “compressor”), and the compressor 1is described later. A refrigerant pipe P is connected to an upper endportion of the compressor 1. A condenser 2, an expansion valve(expansion device) 3, an evaporator 4, and an accumulator 5 aresequentially provided in the refrigerant pipe P. The refrigerant pipe Pis also connected to a side portion of the compressor 1 from theaccumulator 5, thereby forming a refrigeration cycle of therefrigeration cycle apparatus R.

The compressor 1 will be described next. The compressor 1 comprises aclosed vessel 10. A motor portion 11 is accommodated on an upper portionside in the closed vessel 10, and a compression mechanism portion 12 isaccommodated on a lower portion side. The motor portion 11 and thecompression mechanism portion 12 are coupled to each other via arotating shaft 13 interposed therebetween.

A discharge portion la formed by a hole portion is provided in an uppersurface portion of the closed vessel 10, and the refrigerant pipe Pcommunicated with the condenser 2 is connected to the discharge portion1 a. A suction portion 1 b formed by a hole portion is provided in acircumferential wall in a lower portion of the closed vessel 10, and therefrigerant pipe P communicated with the accumulator 5 is connected tothe suction portion 1 b.

The motor portion 11 comprises a rotor 15 and a stator 16. The rotor 15is fitted in and fixed to a rotating shaft 13. An inner circumferentialsurface of the stator 16 faces an outer circumferential surface of therotor 15 with a narrow gap, and the stator 16 is fitted in and fixed toan inner circumferential wall of the closed vessel 10.

The compression mechanism portion 12 will be described below withreference to FIGS. 1 and 2. FIG. 2 is an enlarged longitudinal sectionalview illustrating the compression mechanism portion 12.

The compression mechanism portion 12 comprises a cylinder 20, a mainbearing 21, and a sub-bearing 22. The cylinder 20 is fitted in and fixedto the inner circumferential wall of the closed vessel 10, and aninternal diameter hole S is made in an axial center of the cylinder 20.The main bearing 21 is mounted on an upper surface of the cylinder 20.The sub-bearing 22 is mounted on a lower surface of the cylinder 20. Thecylinder internal diameter hole S is closed by the main bearing 21 andthe sub-bearing 22 to form a space, and the space constitutes acompression chamber (hereinafter referred to as “cylinder chamber”) S.

In the rotating shaft 13, a portion between the motor portion 11 and theupper surface of the cylinder 20 is inserted and journaled in a bearinghole N made in the main bearing 21. In the rotating shaft 13, a portionbetween the lower surface and a lower end of the cylinder 20 is insertedand journaled in a bearing hole N made in the sub-bearing 22.

The main bearing 21 and the sub-bearing 22 comprise flanges 21 a and 22a and cylindrical pivot portions 21 b and 22 b, respectively. Theflanges 21 a and 22 a close the cylinder internal diameter hole S. Thecylindrical pivot portions 21 b and 22 b are projected along axialcenter portions of the flanges 21 a and 22 a while integrated with theflanges 21 a and 22 a, and the cylindrical pivot portions 21 b and 22 bcomprise the bearing holes N in which the rotating shaft 13 isjournaled. Circular grooves K are provided in the main bearing 21 andthe sub-bearing 22, and the circular grooves K are described later.

An eccentric portion 13 a whose center axis is eccentrically located byan eccentric amount e is integrally provided in the rotating shaft 13. Arolling piston (hereinafter simply referred to as “roller”) 25 is fittedin a circumferential surface of the eccentric portion 13 a. The roller25 and the eccentric portion 13 a are accommodated in the cylinderchamber S, and part of an outer circumferential wall of the roller 25 isdesigned to come into linear contact with a circumferential wall of thecylinder chamber S along an axis direction. Accordingly, a position atwhich the outer circumferential wall of the roller 25 comes into contactwith the circumferential wall of the cylinder chamber S moves graduallyin a circumferential direction by the rotation of the rotating shaft 13.

A blade chamber (not illustrated) is provided in the cylinder 20. Acompression spring is accommodated in the blade chamber, and a bladethat receives a back pressure from the compression spring is movablyaccommodated. A leading end edge of the blade is in contact with part ofthe outer circumferential wall of the roller 25 along the axisdirection, and therefore the blade always divides the cylinder chamber Sinto two.

A discharge hole 26 is made in the main bearing 21. A position at whichthe discharge hole 26 is made is located near a region where the bladecomes into contact with the roller 25, and the position constitutes oneside portion of the region. A discharge valve mechanism 27 is providedin the discharge hole 26, and the discharge valve mechanism 27 iscovered with a valve cover 28 mounted on the main bearing 21. A guidehole 28 c is made in the valve cover 28 so as to be opened into theclosed vessel 10.

In the cylinder 20, the hole portion constituting the suction portion 1b is provided across the region where the blade comes into contact withthe roller 25 from the discharge hole 26. The suction portion 1 b iscommunicated with the closed vessel 10 while radially piercing thecylinder 20, and the suction portion 1 b is connected to the refrigerantpipe P communicated with the accumulator 5.

The circular grooves K, provided in the main bearing 21 and thesub-bearing 22, will be described in detail.

The circular groove K provided in the main bearing 21 and the circulargroove K provided in the sub-bearing 22 have the same structure, shape,and dimensions. At this point, only the circular groove K of the mainbearing 21 is described. In the circular groove K of the sub-bearing 22,the same component is designated by the same numeral, and thedescription is not repeated.

The circular groove K is provided from an intersection portion of theflange 21 a and cylindrical pivot portion 21 b constituting the mainbearing 21 to the cylindrical pivot portion 21 b. The circular groove Kcomprises an opening end Kd that is opened to the cylinder chamber S,and the circular groove K is formed deeper from the opening end Kdtoward the side of the motor portion 11 that is the opposite side of thecylinder chamber S.

The opening end Kd of the circular groove K is concentric with thebearing hole N made in the main bearing 21, and the opening end Kd isformed into a ring shape having a predetermined width. In the circulargroove K, a distance between an outer circumferential surface Km and acircumferential surface of the bearing hole N is kept constant from theopening end Kd in a depth direction, while a distance between an innercircumferential surface Kq and the circumferential surface of thebearing hole N increases gradually.

In other words, while the diameter is evenly formed along the axisdirection in the outer circumferential surface Km of the circular grooveK, the inner circumferential surface Kq is tapered such that thediameter increases gradually along the axis direction. Therefore, thewall thickness from the circumferential surface of the bearing hole N tothe inner circumferential surface Kq of the circular groove K becomesminimum (thinnest) at the opening end Kd of the circular groove K andincreases gradually from the opening end Kd in the depth direction.

On the assumption that the inner circumferential surface Kq of thecircular groove K is tapered such that the diameter increases graduallyfrom the opening end Kd that is the side of the cylinder chamber Stoward the opposite side of the cylinder chamber S, a depth L of thecircular groove K is set to at least 40% of a diameter D of the bearinghole N for the later-described reason, where L is a depth of thecircular groove K and D is a diameter (that is also a shaft diameter ofthe rotating shaft 13) of the bearing hole N.

In the circular groove K, because the inner circumferential surface Kqis tapered, a wall thickness b that is a distance between the innercircumferential surface Kq and the circumferential surface of thebearing hole N becomes minimum at the opening end Kd facing the cylinderchamber S. For the later-described reason, the wall thickness b betweenthe inner circumferential surface Kq of the circular groove K and thecircumferential surface of the bearing hole N is set so as to satisfy arelationship of an equation (1):

0.09×diameter D of bearing hole N≧minimum wall thickness b≧0.04×diameterD of bearing hole N   (1)

Assuming that e is an eccentric amount of the eccentric portion 13 a ofthe rotating shaft 13 and r is an outer circumferential radius of theroller 25, the outer circumferential radius g of the circular groove Kis set so as to satisfy relationships of equations (2) and (3) for thelater-described reason:

0.5 mm≦[outer circumferential radius r (mm) of roller 25−eccentricamount e (mm) of eccentric portion 13a]−outer circumferential radius g(mm) of circular groove K   (2)

outer circumferential radius g (mm) of circular groove K>diameter D (mm)of bearing hole N/2+minimum wall thickness b (mm)   (3)

The action of the compressor 1 and freezing action of the refrigerationcycle apparatus R will be described below.

When a current is passed through the motor portion 11 constituting thecompressor 1, the rotor 15 is rotated by a rotating magnetic fieldgenerated by the stator 16, thereby rotating the rotating shaft 13integrated with the rotor 15. A driving torque acts on the rotatingshaft 13 from the motor portion 11, and the eccentric portion 13 aprovided in the rotating shaft 13 is eccentrically rotated whileintegrated with the roller 25 in the cylinder chamber S.

Therefore, a negative pressure is partially generated in the cylinderchamber S, and the refrigerant is introduced from the accumulator 5through the refrigerant pipe P. The refrigerant is introduced into thespace region partitioned by the circumferential surface of the roller25, the circumferential surface of the cylinder chamber S, and theblade, and a volume of the space region is reduced in association withthe eccentric rotation of the roller 25, thereby compressing therefrigerant.

When the space region is minimized, the refrigerant is raised to a hightemperature while attaining a predetermined high-pressure state. Thedischarge valve mechanism 27 is opened by the compressed gasrefrigerant, the refrigerant is introduced to the closed vessel 10through a valve cover 28, and the closed vessel 10 is filled with thegas refrigerant. The high-temperature, high-pressure gas refrigerantwith which the closed vessel 10 is filled is discharged from thedischarge portion 1 a to the refrigerant pipe P.

The condenser 2 performs heat exchange of the gas refrigerant foroutside air or water, and the gas refrigerant is condensed and liquefiedinto a liquid refrigerant. The liquid refrigerant is introduced to theexpansion valve 3 to perform adiabatic expansion, the liquid refrigerantis introduced to the evaporator 4 to perform the heat exchange for airaround a region where the evaporator 4 is disposed, and the liquidrefrigerant is evaporated.

Evaporative latent heat is seized from the surrounding region inassociation with the evaporation of the refrigerant. That is, thefreezing action acts on the surrounding region. The refrigerantevaporated in the evaporator 4 is introduced to the accumulator 5 toperform gas-liquid separation. The refrigerant is sucked into thecylinder chamber S of the compressor 1, the refrigerant is compressedagain to change into the high-temperature, high-pressure gasrefrigerant, and the refrigeration cycle is repeated.

Thus, a suction stroke in which the refrigerant to which the gas-liquidseparation is performed is sucked from the accumulator 5, a compressionstroke in which the sucked refrigerant is compressed, and a dischargestroke in which the compressed refrigerant is discharged arecontinuously performed in the cylinder chamber S constituting thecompression mechanism portion 12.

Particularly, in the compression stroke, the compressive load is appliedto the rotating shaft 13 by the compressed high-pressure gasrefrigerant, whereby the flexural deformation of the rotating shaft 13is generated, from a macroscopic point of view. Specifically, theflexural deformation of the rotating shaft 13 is generated in anopposite direction to the compressive load direction during thecompression action.

However, because the main bearing 21 and the sub-bearing 22 comprise thecircular grooves K set to the above-described conditions, the unevencontact of the rotating shaft 13 with the main bearing 21 andsub-bearing 22 is not generated, irrespective of the flexuraldeformation of the rotating shaft 13, and the smooth rotation issecured.

More specifically, the bearing hole N that is the inner surface of themain bearing 21 is deformed so as to follow the rotating shaft 13 inwhich the flexural deformation is generated by receiving the load, andan area where the evenness of the gap between the rotating shaft 13 andthe main bearing 21 is retained is expanded. Accordingly, the ability toform an oil film of lubrication oil between the rotating shaft 13 andthe main bearing 21 is improved, and the oil film is securely formedeven if the rotating shaft 13 is rotated at low speed.

There are circumstances in which the formation of the oil film canhardly be maintained, such conditions being when the number of rotationsof the rotating shaft 13 is decreased, viscosity of the lubrication oilis reduced, or the compressive load is increased. That is, the contactbetween the rotating shaft 13 and the main bearing 21 makes a transitionto a mixed lubrication state in which not only the rotating shaft 13 andthe main bearing 21 come into contact with each other while the oil filmis interposed therebetween, but also metallic materials come intosolid-state contact with each other due to the surface roughness of therotating shaft 13 and main bearing 21 to support the load.

Even if the solid-state contact cannot be avoided, the surface of thebearing hole N of the main bearing 21 is continuously deformed toprevent the generation of a locally high contact force. The generationof seizing or local bearing abrasion can be prevented to provide thehigh-reliability main bearing 21. Because the sub-bearing 22 comprisesthe circular groove K having completely the same structure, a similareffect can be obtained in the sub-bearing 22.

The circular groove K of the embodiment will be described in comparisonwith a flexible-structure groove described in Jpn. Pat. Appln. KOKAIPublication No. 2004-124834. From the viewpoint of the formation of theoil film, desirably a gap is evenly formed along the axis line directionbetween the main bearing 21 that journals the rotating shaft 13 and therotating shaft 13 in which the flexural deformation is generated byreceiving the compressive load in the cylinder chamber S.

The flexural deformation of the rotating shaft 13 is maximized on theside of the cylinder chamber S in which the compressive load is appliedto the rotating shaft 13 and decreases gradually with distance from theside of the cylinder chamber S. As described above, when the circulargroove K is formed in the main bearing 21, rigidity of the internaldiameter of the main bearing 21 is low on the side of the cylinderchamber S in which the rotating shaft 13 has the large flexuraldeformation, and the rigidity increases gradually with distance from theside of the cylinder chamber S.

Therefore, the inner surface of the main bearing 21 is deformed byfollowing the deformation of the rotating shaft 13, and the deformablecircular groove K is formed deeper than the flexible-structure groove,so that the circular groove K is greatly deformed in a wide area tofollow the rotating shaft 13. Additionally, the rigidity of the internaldiameter of the main bearing 21 increases gradually with distance fromthe side of the cylinder chamber. S, so that a fluctuation in loadapplied to the main bearing 21 in the axis direction can be reduced.

On the other hand, in the flexible-structure groove, because the wallthickness between the groove inner surface and the circumferentialsurface of the bearing hole is kept constant over the total length ofthe groove, the rigidity of the circumferential surface of the bearinghole is kept constant. Therefore, the rigidity is small in the grooveportion, the rigidity increases rapidly in the portion in which thegroove is terminated, and the fluctuation in load applied to the bearingalso increases. Accordingly, the oil film is easily broken in theportion in which the groove is terminated. This cannot be solved even ifthe groove depth is simply increased.

In the embodiment, the circular groove K is provided, and the depth ofthe groove K and the wall thickness between the groove K and the bearinghole N are increase to enhance the strength. The rigidity of theinternal diameter of the main bearing 21 increases with distance fromthe side of the cylinder chamber S, the oil film is evenly formed in thewhole of the main bearing 21, and the fluid lubrication state can bemaintained in the wide operating area.

Even if the contact between the rotating shaft 13 and the main bearing21 makes the transition from the fluid lubrication state to the mixedlubrication state in which the lubrication state including thesolid-state contact state is maintained, because the circular groove Kis deep and flexible, the solid-state contact is generated in the depthrange of the circular groove K in which the elastic deformation can begenerated, and the main bearing 21 is elastically deformed to preventthe uneven contact with the rotating shaft 13. Therefore, seizing andthe like are not generated.

As described above, there is the setting condition that the innercircumferential surface Kq of the circular groove K is tapered. Thesetting condition is fixed on the following basis. First, the basis onwhich the depth L of the circular groove K is set to at least 40% of thediameter D of the bearing hole N will be described, on the assumptionthat the inner circumferential surface Kq of the circular groove K istapered such that the diameter increases gradually from the surfacefacing the cylinder chamber S toward the opposite side of the cylinderchamber S.

In the bearing hole N of the main bearing 21, the portion in which thecircumferential surface of the rotating shaft 13 is particularlyeffectively journaled is a portion from an end portion of the bearinghole N to a length corresponding to the diameter of the bearing hole N.The depth L of the circular groove K is formed equal to or more than 40%of the diameter D of the bearing hole N.

Therefore, the inner surface (bearing hole N) of the main bearing 21 isdeformed so as to follow the deformation of the rotating shaft 13, whichdesirably affects the formation of the oil film between the rotatingshaft 13 and the main bearing 21 and the contact of the rotating shaft13 with the main bearing 21 due to the deformation of the rotating shaft13.

This can be described with reference to FIG. 6. FIG. 6 is acharacteristic diagram illustrating a groove depth effect. In FIG. 6, ahorizontal axis indicates the depth of the circular groove K, and avertical axis indicates the oil film thickness of the lubricant oilformed between the rotating shaft 13 and the main bearing 21 and thecontact force between the rotating shaft 13 and the main bearing 21. InFIG. 6, a solid-line indicates the contact force and a broken-lineindicates the oil film thickness. Where the depth of the circular grooveK is indicated by a ratio to the shaft diameter (diameter) D of therotating shaft 13 (bearing hole N).

When the depth of the circular groove K in which the innercircumferential surface Kg is tapered becomes zero, the contact forcebetween the rotating shaft 13 and the main bearing 21 becomes maximum(100), and the oil film is hardly formed. The oil film is formed in thethinnest state at a point where the contact force is weakened to someextent. The contact force decreases rapidly with increasing depth of thecircular groove K, and the oil film thickness is thickened in inverseproportion to the decreasing contact force.

Particularly, when the depth of the circular groove K exceeds 0.4 (40%of the shaft diameter ratio), a degree to which the contact forcedecreases changes from the rapidly decreasing state to the graduallydecreasing state, the oil film thickness exceeds a necessary oil filmthickness (1), and the oil film thickness is maintained at 1 or more.

In the fluid lubrication “state in which only the oil film of thelubrication oil is interposed between the rotating shaft 13 and the mainbearing 21, the oil film thickness is thickened by increasing the groovedepth, a tilt of the rotating shaft 13 increases to keep the oil filmthickness substantially constant when the depth of the circular groove Kbecomes at least 40% of the shaft diameter ratio of the rotating shaft13.

On the other hand, the contact load of the rotating shaft 13 and themain bearing 21 in the mixed lubrication state exhibits a characteristicin which the contact load can be reduced with increasing depth of thecircular groove K. However, when the depth of the circular groove Kbecomes at least 40% of the shaft diameter ratio of the rotating shaft13, the tilt of the rotating shaft 13 increases, and a decreasing ratioof the contact load becomes small.

In the circular groove K whose inner circumferential surface Kq istapered, the wall thickness b that is the distance between the innercircumferential surface Kq and the bearing hole N becomes minimum(thinnest) at the opening end Kd facing the cylinder chamber S.

The minimum wall thickness b between the inner circumferential surfaceKg of the circular groove K and the circumferential surface of thebearing hole N is set so as to satisfy the relationship of the equation(1):

0.09×diameter D of bearing hole N≧minimum wall thickness b≧0.04×diameterD of bearing hole N   (1)

This can be described with reference to FIG. 7. FIG. 7 is acharacteristic diagram illustrating a circular groove minimum wallthickness effect. In FIG. 7, the horizontal axis indicates the minimumwall thickness (shaft diameter ratio) b of the circular groove K, andthe vertical axis indicates the contact force. In FIG. 7, the solid-linechange indicates the contact force, and a maximum allowable contactforce is set to 0.5.

When the minimum wall thickness b of the circular groove K decreasesexcessively, a lack of rigidity is generated in the main bearing 21, andthe deformation becomes large. At this point, even if the oil filmthickness can be secured in the fluid lubrication state, the contactload increases in the mixed lubrication state.

On the other hand, when the minimum wall thickness b of the circulargroove K increases excessively, the rigidity increases excessively tohardly generate the deformation, and the contact load also increases inthe mixed lubrication state. Therefore, the proper value of the minimumwall thickness to the contact load is set as illustrated in FIG. 7 andthe equation (1).

Assuming that e is the eccentric amount of the eccentric portion 13 athat is provided integral with the rotating shaft 13 and r is the outercircumferential radius of the roller 25, the outer circumferentialradius g of the circular groove K is set so as to satisfy therelationships of the equations (2) and (3):

0.5 mm≦[outer circumferential radius r (mm) of roller 25−eccentricamount e (mm) of eccentric portion 13a]−outer circumferential radius g(mm) of circular groove K   (2)

outer circumferential radius g (mm) of circular groove K>diameter D (mm)of bearing roller N/2+minimum wall thickness b (mm)   (3)

When the opening end Kd of the circular groove K is communicated withthe cylinder chamber S, the refrigerant introduced to the cylinderchamber S remains partially in the circular groove K, and the circulargroove K becomes a dead volume. Therefore, in order to prevent the deadvolume of the circular groove K, a minimum seal width is formed to exerta seal function between an external diameter of the roller 25 and anexternal diameter of circular groove K.

Particularly, the equation (2) can be described with reference to FIG.8. FIG. 8 illustrates a minimum seal width effect. In FIG. 8, thehorizontal axis indicates a minimum seal width (mm), and the verticalaxis indicates a performance ratio.

The performance ratio is 0.2 when the minimum seal width becomes 0, andthe performance ratio does not change even if the minimum seal widthincreases to about 0.3 mm. The performance ratio increases when theminimum seal width exceeds about 0.3 mm, and the performance ratioincreases rapidly when the minimum seal width exceeds 0.4 mm.

The performance ratio becomes a peak when the minimum seal width isabout 0.5 mm, and the performance ratio is substantially kept constanteven if the minimum seal width increases from about 0.5 mm.

In the equation (2), [outer circumferential radius r (mm) of roller25−eccentric amount e (mm) of eccentric portion 13 a]−outercircumferential radius g (mm) of circular groove K is the minimum sealwidth. As can be seen from FIG. 8, the minimum seal width of 0.5 mm ormore is required.

As described above, the inner circumferential surface Kq of the circulargroove K is tapered, and setting of a slope angle θ becomes one ofnecessary conditions. That is, the contact force between the rotatingshaft 13 and the main bearing 21 varies depending on the slope angle θ.The circular groove K is formed such that the slope of the innercircumferential surface Kq increases (the slope angle θ decreases) asmuch as possible, thereby exerting the large contact load reducingeffect.

FIG. 9 is a characteristic diagram illustrating a groove slope effect.In FIG. 9, the horizontal axis indicates the slope of the innercircumferential surface Kg of the circular groove K, and the verticalaxis indicates the contact force between the rotating shaft 13 and themain bearing 21.

The contact force is maximized (1 or more) when the slope of thecircular groove K is close to zero (0). With increasing groove slope,the contact force decreases, and therefore the oil film thicknessincreases as described above.

Further, as illustrated in FIG. 2, there is another setting conditionthat the main bearing 21 comprises the flange 21 a whose wall thicknessH is set to the depth L of the circular groove K or less.

Therefore, the rigidity of the coupling portion between the cylindricalpivot portion 21 b and the flange 21 a that supports the whole of themain bearing 21 is reduced to deform the whole of the main bearing 21,whereby a property of following the rotating shaft 13 is enhanced toimprove the effect of the circular groove K.

FIG. 3 is an enlarged longitudinal sectional view of a compressionmechanism portion 12 according to a second embodiment of the invention.Because a basic configuration of a compression mechanism portion 12 isidentical to that of FIG. 2, the same component is designated by thesame numeral (only main part), and the description of the same componentis not repeated. In the second embodiment, a diameter D1 of a portion(bearing hole Na) that is journaled in a main bearing 21 of a rotatingshaft 13 differs from a diameter D2 of a portion (bearing hole Nb) thatis journaled in a sub-bearing 22. Actually, the diameter D1 of theportion journaled in the main bearing 21 of the rotating shaft 13 isformed larger than the diameter D2 of the portion journaled in thesub-bearing 22 (D1>D2).

Because the diameter D1 is formed larger than the diameter D2, it isnecessary to secure a seal width of a circular groove K with respect toa cylinder chamber S in an end face of a roller 25. Therefore, an innercircumferential surface Kq of the circular groove K is hardly tapered,and a groove Ka having an even width in the depth direction is provided.

That is, the tapered inner circumferential surface Kq of the circulargroove K is provided only in a rotating shaft portion that is journaledin the sub-bearing 22 having a small diameter, and the seal width of theend face of the roller 25 is secured with respect to the cylinderchamber S.

Because the length in the axis direction of the cylindrical pivotportion 22 b is shorter than that of the main bearing 21, the flexuraldeformation becomes large, and the load also becomes large. Therefore,the circular groove K whose inner circumferential surface Kq is taperedis extremely advantageously provided.

In the circular groove K whose inner circumferential surface Kq istapered, the dimensions and configuration are similar to those of thefirst embodiment, and the effect similar to that of the first embodimentis obtained. However, the overlapping description is not repeated.

FIG. 4 is a longitudinal sectional view illustrating a hermetic typecompressor 1A according to a third embodiment of the invention with partof the hermetic type compressor 1A omitted.

Basically, the configuration in which a motor portion 11 and acompression mechanism portion 12A that is coupled to the motor portion11 with a rotating shaft 13 interposed therebetween are accommodated ina closed vessel 10 is similar to that of the first embodiment.

The compression mechanism portion 12A is a two-cylinder type compressor1A that comprises two cylinders 20A and 20B that are provided above andbelow an intermediate partition plate 30. Each of the cylinders 20A and20B comprises an internal diameter hole Sa. The internal diameter holeSa of the cylinder 20A on the upper side is closed by a main bearing 21and the intermediate partition plate 30 to form the first cylinderchamber Sa.

The internal diameter hole Sb of the cylinder 20B on the lower side isclosed by a sub-bearing 22 and the intermediate partition plate 30.Eccentric portions 13 a and 13 b and a roller 25 are accommodated in thefirst cylinder chamber Sa and the second cylinder chamber Sb,respectively. The eccentric portions 13 a and 13 b are provided whileintegrated with the rotating shaft 13, and the eccentric portions 13 aand 13 b have a phase difference of 180°. The roller 25 is fitted in theeccentric portions 13 a and 13 b.

A diameter of a portion journaled in the main bearing 21 of the rotatingshaft 13 is equal to a diameter of a portion journaled in thesub-bearing 22. In other words, diameters of bearing holes N made in themain bearing 21 and sub-bearing 22 are equal to each other.

Circular grooves K opened to the cylinder chambers Sa and Sb areprovided in the main bearing 21 and the sub-bearing 22. An innerperipheral surface of the circular groove K is tapered such that adiameter of the inner peripheral surface increases gradually from thesurface facing each of the cylinder chambers Sa and Sb toward theopposite side of the cylinder chamber. The depth of the circular grooveK is set to at least 40% of the diameter of the bearing hole.

Because all the above-described setting conditions are included in thehermetic type compressor 1A of the third embodiments, similar effectsare obtained in both the main bearing 21 and the sub-bearing 22.

FIG. 5 is a longitudinal sectional view illustrating a hermetic typecompressor 1B according to a fourth embodiment of the invention withpart of the hermetic type compressor 1A omitted.

Basically, the hermetic type compressor 1B of the fourth embodimentcomprises a compression mechanism portion 12B having a configurationsimilar to that of the two-cylinder type compression mechanism portion12A of the third embodiment (see FIG. 4).

In the fourth embodiment, a diameter D1 of a portion journaled in a mainbearing 21 of a rotating shaft 13 differs from a diameter D2 of aportion journaled in a sub-bearing 22. The diameter D1 of the portionjournaled in the main bearing 21 of the rotating shaft 13 is formedlarger than the diameter D2 of the portion journaled in the sub-bearing22 (D1>D2).

Accordingly, in the compression mechanism portion 12B, similarly to thecompression mechanism portion 12 of the second embodiment (see FIG. 3),because the diameter D1 is formed larger than the diameter D2, it isnecessary to secure a seal width of a circular groove K with respect toa cylinder chamber S in an end face of a roller 25. Therefore, an innercircumferential surface Kg of the circular groove K is hardly tapered,and a groove Ka having an even width in the depth direction is provided.

The tapered inner circumferential surface Kq of the circular groove K isprovided only in a portion of a rotating shaft 13 that is journaled inthe sub-bearing 22 having a small diameter, and the seal width of theend face of the roller 25 is secured with respect to the cylinderchamber S.

Because the length in the axis direction of the cylindrical pivotportion 22 b is shorter than that of the main bearing 21, the flexuraldeformation becomes large, and the load also becomes large. Therefore,the circular groove K whose inner circumferential surface Kq is taperedis extremely advantageously provided.

FIG. 10 is a longitudinal sectional view of the hermetic type compressor1A according to a modification of the third embodiment of the invention,and the refrigeration cycle is omitted in FIG. 10.

Basically, the hermetic type compressor 1A of the modification of thethird embodiment comprises the two-cylinder type compression mechanismportion 12A of the third embodiment (see FIG. 4), diameters of bearingholes N made in a main bearing 21 and a sub-bearing 22 are equal to eachother, and the main bearing 21 and the sub-bearing 22 comprise circulargrooves K.

In the modification of the third embodiment, a discharge valve mechanism27 for a first cylinder chamber Sa is provided in the main bearing 21, adischarge valve mechanism 27 for a second cylinder chamber Sb isprovided in the sub-bearing 22, and a discharge valve mechanism 27A forthe first cylinder chamber Sa and a discharge valve mechanism 27A forthe second cylinder chamber Sb are provided in an intermediate partitionplate 30A that is interposed between two cylinders 20A and 20B.

Because the intermediate partition plate 30A comprises the two dischargevalve mechanisms 27A, the intermediate partition plate 30A is dividedinto two in a thickness direction. As described later, the two dischargevalve mechanisms 27A of the intermediate partition plate 30A are mountedwhile overlapping each other when viewed from above.

FIG. 11 is a plan view of the intermediate partition plate 30A whenviewed from a side of a surface in which the discharge valve mechanisms27A overlap each other.

As illustrated by a solid-line arrow of FIG. 11, a gas refrigerant thatis discharged from a discharge holes 26 made in each of the dividedintermediate partition plates 30A is guided to the outside from acommunication hole 32 through a groove 31 provided in each of theintermediate partition plates 30A.

FIG. 12 is a longitudinal sectional view of a region where the dischargevalve mechanisms 27A are provided in the intermediate partition plates30A divided into two.

The discharge valve mechanism 27A comprises a discharge valve 33 and adischarge valve guard 34 a. One end of the discharge valve 33 issupported while separated from a discharge hole 26. The discharge valve33 is formed by a thin spring plate, and the other end of the dischargevalve 33 is in close contact with the discharge hole 26 so as to closethe discharge hole 26. The discharge valve guard 34 a is formed by athick plate piece having rigidity, and the discharge valve guard 34 a isgently bent from a support portion at one end toward the discharge hole26 at the other end.

A pressure at each of cylinder chambers Sa and Sb increases by thecompression action of the refrigerant, the discharge valve 33 is pressedwhen the pressure reaches a predetermined value, and the discharge valve33 is elastically deformed to open the discharge hole 26. Accordingly,the high-pressure gas refrigerant compressed by each of the cylinderchambers Sa and Sb is discharged from the discharge hole 26. Thedischarge valve guard 34 a receives the elastically-deformed dischargevalve 33 to regulate further deformation, thereby preventing metalfatigue of the discharge valve 33 as much as possible.

The discharge valve guard 34 a has a specific thickness because thedischarge valve guard 34 a has the necessary rigidity. One end of thedischarge valve guard 34 a mounted on the intermediate partition plate30A is formed into a flat shape, and the discharge valve guard 34 a isbent into a predetermined curved shape from the flat-shape leading endto the other end facing the discharge hole 26. Therefore, the leadingend of the discharge valve guard 34 a is formed at a certain level froma flat surface formed in the mounting portion.

When the intermediate partition plate 30A directly comprises thedischarge valve mechanism 27A, the wall thickness of the intermediatepartition plate 30A increases considerably, and the compressionmechanism portion 12A is lengthened in the axis direction, which leadsto enlargement of the compressor 1A.

When the intermediate partition plate 30A is thickened, an intervalbetween the first cylinder chamber Sa and the second cylinder chamber Sbis lengthened, and the distance between the eccentric portions 13 a ofthe rotating shafts 13 that are accommodated in the first cylinderchamber Sa and the second cylinder chamber Sb. This leads to thedegradation of the rigidity of the rotating shaft 13 to cause theincrease of flexural deformation, amplification of wobbling, and thedegradation of the reliability.

Therefore, as illustrated in a first example of FIG. 12, in thedischarge valve guard 34 a, a flat portion mounted on the intermediatepartition plate 30A has the same wall thickness, and a bent portion Ufacing the discharge hole 26 is tapered such that a wall thicknessdecreases gradually toward the leading end and such that the wallthickness in section becomes the thinnest in the leading end portion.

Because the discharge valve guard 34 a receives the force of thedischarge valve 33, strength is required for the discharge valve guard34 a, and the discharge valve guard 34 a is formed with a predeterminedthickness. However, a stress is not applied to the leading end of thebent portion U too much, and no problem occurs even if a section of theleading end of the bent portion U is thinned into the tapered shape.

Therefore, a height of the discharge valve guard 34 a can be reduced todecrease the wall thickness of the intermediate partition plate 30A. Asthe height of the compression mechanism portion 12A is reduced, thedistance between the eccentric portions 13 a of the rotating shafts 13can be shortened to reduce the flexural deformation or wobbling of therotating shaft 13, thereby improving the reliability.

Meanwhile, the discharge valve mechanisms 27 of the main bearing 21 andsub-bearing 22 are removed, the discharge valve mechanism 27A for thefirst cylinder chamber Sa and the discharge valve mechanism 27A for thesecond cylinder chamber Sb may be provided only in the intermediatepartition plate 30A.

Alternatively, as illustrated in a second example of FIG. 13, only aleading end Z of the bent portion is processed, although the platethickness is evenly formed from the mounting portion to the bent portionwithout changing a configuration of the discharge valve guard 34 a.

That is, at the leading ends Z of the discharge valve guards 34 a,surfaces facing each other that are the surfaces that do not collidewith the discharge valves 33 are cut into flat shapes so as to beparallel. Therefore, the distance between the mounting portions of thetwo discharge valve guards 34 a can further be shortened to minimize thethickness of the intermediate partition plate 30A, so that theabove-described effect is obtained.

The invention is not limited to the embodiments, but variousmodifications can be made at an implementation stage without departingfrom the scope of the invention. Various inventions can be made byappropriately combining a plurality of constituents disclosed in theembodiments.

According to the invention, according to the flexural deformation of therotating shaft due to the compressive load in the cylinder, the unevencontact with the rotating shaft is prevented in at least one of the mainbearing and sub-bearing, thereby achieving the enhancement of thereliability and the longer operation life. Additionally, the hermetictype compressor is provided to form the refrigeration cycle, therebyimproving refrigeration efficiency.

1. A hermetic type compressor in which a motor portion and a compressionmechanism portion that are coupled to the motor portion with a rotatingshaft interposed therebetween are accommodated in a closed vessel,wherein the compression mechanism portion comprises: a cylinder thatcomprises an internal diameter hole; and a main bearing and asub-bearing, in which a bearing hole that journals the rotating shaft isprovided and the internal diameter hole of the cylinder is closed toform a compression chamber in the compression mechanism portion, atleast one of the main bearing and the sub-bearing has a circular groovethat is opened toward the compression chamber side, an innercircumferential surface of the circular groove is tapered such that adiameter increases gradually from the compression chamber side toward anopposite side of the compression chamber side, and a depth L of thecircular groove is set to at least 40% of a diameter D of the bearinghole.
 2. The hermetic type compressor according to claim 1, wherein, inthe main bearing or sub-bearing that comprises the circular groove, aminimum wall thickness b between an inner circumferential surface of thecircular groove and a circumferential surface of the bearing hole is setso as to satisfy a relationship of an equation (1):0.09×diameter D of bearing hole≧minimum wall thickness b≧0.04×diameter Dof bearing hole   (1).
 3. The hermetic type compressor according toclaim 1, wherein a compression chamber of the compression mechanismportion accommodates an eccentric portion that is eccentrically providedwhile integrated with the rotating shaft and a rolling piston that isfitted in the eccentric portion to rotate eccentrically in thecompression chamber in association with rotation of the rotating shaft,and assuming that e is an eccentric amount of the eccentric portion andr is an outer circumferential radius of the rolling piston, an outercircumferential radius g of the circular groove satisfies relationshipsof equations (2) and (3):0.5 mm≦[outer circumferential radius r (mm) of rolling piston−eccentricamount e (mm) of eccentric portion]−outer circumferential radius g (mm)of circular groove   (2)outer circumferential radius g (mm) of circular groove>diameter D (mm)of bearing hole/2+minimum wall thickness b (mm)   (3).
 4. The hermetictype compressor according to claim 1, wherein the main bearing and thesub-bearing have a flange whose wall thickness is set to a depth L ofthe circular groove or less.
 5. A refrigeration cycle apparatuscomprising: the hermetic type compressor according to claim 1; acondenser; an expansion device; and an evaporator.